1. Field of Invention
This invention relates to an air volume regulator, more specifically to an improved “airflow powered” air volume regulator as used to control the flow of conditioned air in the ductwork of heating, ventilating and air conditioning (HVAC) systems, clean rooms or fume hood systems.
2. Background of the Invention
In HVAC systems, air is supplied from a central air conditioning system to several outlet devices such as grilles and diffusers in the rooms or spaces being conditioned. Once a HVAC system is installed, the airflow through the ductwork system must be adjusted or balanced. This insures that each room or space obtains the specified volumes of conditioned air from the central system. In its simplest form, this can be done by using manually adjustable dampers. They are placed within the supply air and return air ductwork to reduce the airflow in areas where it exceeds the specified amount. There is an inherent problem with this method. When one damper is adjusted, the pressure level throughout the ductwork system will change. Any change in the ductwork system pressure will affect the flow or air past every other damper including the previously adjusted dampers. On large systems, it quickly becomes impractical to attempt to balance the ductwork system using dampers alone. To solve this problem, air volume regulators are added. They are designed to limit the supply of conditioned air to the desired amount and this, irregardless of the pressure at their inlet. Also, once calibrated, the airflow is not affected by subsequent variations in system pressure. The accepted industry standard airflow variation is +/−5% of the specified airflow volume (4.7 L/s below 94 L/s or 10 cfm below 200 cfm) over the airflow regulator's pressure range.
Furthermore, heating, cooling and ventilating loads in a room or space vary in time. It has become common practice to stabilize the temperature of the rooms or spaces by:                (a) varying the volume of conditioned air supplied to each room or space or        (b) using heating coils downstream from the air volume regulator to heat the volume of cool conditioned air being supplied to the room or space or        (c) using a dual ductwork layout for the HVAC system known as dual-duct system: one supplying hot air, the second supplying cool air and a mixing valve upstream of the air volume regulator or        (d) using combinations of the above.        
Air volume regulators fall into one of two general groups based on the source of energy that is used to drive them: “airflow powered” and “externally powered”.
“Airflow powered” air volume regulators function using the energy of the air flowing in the ductwork system. This source of energy is in the form of air static pressure and air velocity pressure (called dynamic pressure). The scope of this invention is limited to improved “airflow powered” air volume regulators.
Briefly, “externally powered” air volume regulators operate using an external energy source such as pneumatic pressure or electricity. They require an airflow sensor, a signal amplifier, an actuator and an adjustable airflow restricting device or damper to regulate the flow of air.
As mentioned above, the energy source used to drive the “airflow powered” air volume regulator comes from two types of pressure present in HVAC systems: static pressure and dynamic pressure. The static pressure induces the movement of the air through the ductwork towards the outlets while the dynamic or velocity pressure is generated by the movement of the air at a given point within the ductwork. The higher the air velocity the greater the dynamic pressure.
The flow of air through the ductwork of an HVAC system is governed by the following basic formulas:
No. 1: The relationship between air velocity and dynamic pressure is given by the followingPd=Dynamic Pressure=Constant×(Va)2 
No. 2: The sum of the static pressure and dynamic pressure is called the total pressure:Pt=Ps+Pd 
No. 3: The airflow rate is equal to the air velocity times the area of the duct cross section.Qa=Va×Aduct                 where Qa=Air volume rate        Va=air velocity        Aduct=area of the duct cross section        
A conclusion of formula no. 2 is that, under the idealized conditions of constant total pressure (i.e. no losses due to friction and turbulence), the static pressure and dynamic pressure can be converted from one to the other. A decrease of one entails an equal increase of the other.
A conclusion of formula no. 3 is that, for a constant volume flow, an increase in the duct cross section entails a proportional decrease in the air velocity. Conversely, a decrease in the duct cross-section entails a proportional increase in the air velocity.
Combining the 3 formulas, we can conclude that, for a constant total pressure (i.e. no losses due to friction and turbulence):                an increase in the duct cross section means                    a decrease in dynamic pressure (associated with a decrease in air velocity)            an equal increase in static pressure. This is known as “static regain”.                        a decrease in the duct cross section means                    an increase in dynamic pressure (again associated to the air velocity) and            an equal decrease in static pressure.                        
3. Background of the Invention—Discussion of Prior Art
A—INTERACTION WITH THE HVAC SYSTEM
As stated above, an air volume regulator is required when the airflow rate in the ductwork system exceed the desired amount. This happens when more static pressure is present in the air duct than is required to move the air to the outlet devices. Although all air volume regulators inherently create a pressure loss due to air friction and air turbulence as the air flows through them, prior art “airflow powered” air volume regulator also requires some amount of pressure to drive the flow control means. The sum of these pressure losses is called the regulator minimum static pressure. Under most conditions, this extra control pressure required to drive the flow control means is of little consequence. Excess pressure is usually present at the inlet of the air volume regulator. In the cases where the air volume regulator is installed in a area with little or no excess static pressure (i.e. at the far limits of the central air distribution system), the control pressure requirement may be greater than the available excess pressure. Consequently, the air volume regulator will be incapable of controlling the airflow and the desired airflow rate will not be attained. Thus a desirable characteristics of an air volume regulator is that it have a negligible pressure loss due to air friction and turbulence and more important still, require virtually no static pressure to bring the flow of air under control. Although many forms of prior art have been proposed, none have met this challenge.
B—OPERATION
An air volume regulator functions by varying its internal airflow passageway(s) so as to maintain a substantially constant airflow rate. In “airflow powered” air volume regulators, the air pressure and velocity drive some form of airflow restricting means. A spring is then used to counterbalance the forces acting on the restricting means (called the counterbalance spring).
The graph in FIG. 1 shows, for a constant flow of air, the variation in the cross sectional area of the narrowest portion or throat of the airflow passageway as the static pressure differential across the air volume regulator increases from 25 pascals to 1000 pascals (0.1″ to 4″ w.g).
The graph in FIG. 1 is derived from the following:                (a) with the throat open to its maximum, the static pressure drop is 25 pascals (0.1″ w.g.); this pressure is mainly to drive the flow control means and frictional loses due to turbulence,        (b) the variation in the area of the throat is inversely proportional to the air velocity at the throat, i.e. if the velocity goes up, the area must go down proportionally, a corollary of formula 3 above,        (c) the static pressure differential or static pressure drop between the inlet and outlet of the air passageway is substantially converted to dynamic pressure in the throat of the passageway. This assumes that the total pressure remains substantially constant (negligible loses due to friction or turbulence)        (d) in the throat, the static pressure drop is proportional to the square of velocity at that point (the drop in static pressure is totally converted to dynamic pressure) i.e. if the static pressure drop doubles, the velocity in the throat will quadruple.        
Combining the three previous statements, an exponential equation is obtained:
      Area    =            constant                      static  pressure  drop                            where the constant depends on the air volume regulator design.        
FIG. 1 is a graph of this exponential equation. The cross-sectional area of the throat must drop quickly as the pressure differential rises. It then starts to level off to a point were very little reduction of the throat is required to control the airflow (approximately 500 pascals or 2″ w.g.). Also, as the pressure differential rises from 25 pascals to 125 pascals (0.1″ to 0.5″ w.g.), the area of the throat must be reduced by over 50%. Furthermore, since the air volume regulator contains moving parts, mechanical friction is present and this inhibits the reduction of the throat. Thus to operate reliably, the airflow regulator must be capable of initiating and maintaining control of the airflow using and/or amplifying the very low forces generated by these low pressures. I have found no prior art that proposes an “airflow powered” air volume regulator with this latter capability.
The sum of all mechanical friction in an airflow regulator generates an adverse effect on its operation. This can be visualized by drawing an hysteresis graph: a graph is plotted of the airflow rate versus inlet pressure as the pressure is slowly increased up to the upper limit of the airflow regulator's pressure range then, on the same graph, is plotted the airflow rate versus inlet pressure as the pressure is slowly decreased down to the lower limit of the pressure range. The two curves do not coincide with one another. The reason for this is that as the pressure increases, the air velocity increases and the airflow regulator tends to reduce the airflow passageways so as to maintain the specified airflow rate. But mechanical friction within it tends to resist this reduction and the correct passageway size is not attained. The airflow passageways are consequently a little to wide and thus the airflow rate will be slightly above the specified airflow rate. Conversely, as the pressure decreases from the upper limit of the pressure range, the air velocity decreases and the airflow regulator tends to increase the airflow passageways so as to maintain the desired airflow rate. Again mechanical friction within it tends to resist this increase. The airflow passageways are consequently a little to narrow and thus the airflow rate will be slightly below the specified airflow rate. The more mechanical friction is found in an airflow regulator, the greater the difference between the specified airflow rate and the actual airflow rate. As outlined above, this difference must not exceed 5% of the specified airflow rate.
In practice, attempts made in commercially available prior art to reduce the regulator minimum static pressure below 100 pascals (0.3″ w.g.) for “airflow powered” units have been unsuccessful. In referring to U.S. Patents such as 3,049,146 to Hayes (1962), 2,890,716 to Werder (1959), 3,338,265 to Kennedy (1967) and 4,009,826 to Walker (1977), forcing the airflow to pass through perforated screens has a particularly adverse effect on the regulator minimum static pressure (this generates high airflow friction and turbulence losses). As mentioned above, particular attention must be paid to mechanical friction in the moving parts of the air volume regulator. All prior art embodiments using components sliding on shafts encounter, over time, binding of some kind when dirt particles get lodged in the sliding bearing. This is the case with U.S. Pat. Nos. 3,204,664 to Gorchev et al. (1965), 3,763,884 to Grassi et al. (1973), 3,958,605 to Nishizu et al. (1976) and 4,009,826 to Walker (1977). Adding lubricant does not improve this inconvenience since again, over time, the lubricant attracts dirt particles to form a abrasive paste that resists movement.
C—COUNTERBALANCE SPRING
The air volume regulator should include a means to easily vary the desired airflow rate's setpoint at will once installed in the ductwork system. This is made necessary by changing conditions such as occupancy of the areas, insolation and outside temperature. It is common practice to add an optional actuator to adjust the setpoint mechanism and control it remotely. As noted above, a spring is used to counterbalance the forces acting on the flow restricting means. As stated in U.S. Pat. Nos. 4,306,585 to Manos (1981), 4,009,826 to Walker (1977), 3,958,605 to Nishizu (1976), 3,942,552 and 3,939,868 to Logsdon (1976), 3,763,884 to Grassi et al. (1973), 3,565,105 to Murakami (1971) and 3,037,528 to Baars et al. (1962), varying the initial load of the counterbalance spring by adjusting the initial spring deflection is only effective for small variations in the airflow setpoint. As stated in these patents, the spring quickly become either to stiff or to soft and the air volume regulator ceases to adequately control the airflow. Although not stated in the following patents, I also believe this to be true for U.S. Pat. Nos. 4,633,900 to Suzuki (1987), 3,967,642 to Logsdon (1976), 3,433,410 to Warren (1969), the embodiment in FIG. 7 of 3,276,480 to Kennedy (1966), 3,204,664 to Gorchev et al. (1965), 3,049,146 to Hayes (1962) and 2,890,716 to Werder (1959). Thus for a particular air volume regulator at a given flow rate corresponds a spring stiffness known as its spring rate or spring constant and is defined as the force generated divided by the spring deflection. In general terms, the spring rate is the “force-displacement” characteristic of the spring. To vary the airflow setpoint requires that the spring stiffness be varied or several springs be used over the operating range of the regulator. As shown in U.S. Pat. Nos. 4,009,826 to Walker (1977), 3,939,868 (1976) and 3,942,552 (1976) to Logsdon and my own U.S. Pat. No. 4,130,132 (1978), relatively complex mechanisms are proposed to vary the spring stiffness.
D—ACTUATORS
As mentioned above, an actuator may be used to action the flow rate setpoint mechanism. The control signal to the actuator is generally supplied by a thermostat. Actuators may be either pneumatic or electric driven. But pneumatic actuators can be a problem when the flow restricting forces applied to the counterbalance spring are also carried by the actuator. Pneumatic actuators have an inherent load dependant stroke or travel due to the compressibility of the air pushing the actuator's piston. Since the flow restricting forces vary in time due to changes in the static pressure upstream from the air volume regulator, so does the load on the counterbalance spring and thus the actuator. The pneumatic actuator's piston will move or slip under the varying load with the ensuing unjustified change in the flow rate setpoint. This phenomena is clearly outlined in the report “Factors that work to defeat the application of the “spring and cone” type valves in laboratory and other precision airflow systems” by Swiki A. Anderson, Ph.D., P. E., (Swiki Anderson & Associates, Inc. 1516 Shiloh Avenue, Bryan, Tex. 77803). The thermostat will sense a variation in the temperature of the room caused by the change in the flow rate and adjust the pressure to the pneumatic actuator to rectify the unjustified change and its ensuing discomfort to the occupants. This is the case in U.S. Pat. Nos. 4,633,900 to Suzuki (1987), 4,175,583 to Finkelstein et al. (1979), 3,958,605 to Nishizu et al. (1976), 3,942,552 to Logsdon (1976), 3,204,664 to Gorchev (1965) and my own U.S. Pat. No. 4,130,132 (1978).
Further concerns involving the actuator are:                To facilitate field servicing and repairs, the actuator should not be situated inside the air volume regulator or its housing such as U.S. Pat. Nos. 3,976,244, 3,942,552 and 3,939,868 to Logsdon (1976) and my own U.S. Pat. No. 4,130,132 (1978)        Its replacement should not affect the calibration of the air volume regulator such as my own U.S. Pat. No. 4,130,132 (1978).E—ZERO FLOW        
When an actuator is used to vary the flow rate set point and under certain conditions, it is common practice in HVAC systems to restrict the flow completely (substantially zero flow is the accepted industry standard leakage of 2% of the maximum airflow capacity at the maximum operating pressure of the regulator). The flow restricting means must then be able to block the flow of air through the air volume regulator. In prior art, this is not possible with U.S. Pat. No. 3,958,605 to Nishizu et al. (1976), U.S. Pat. No. 4,009,826 to Walker (1977) or U.S. Pat. No. 4,633,900 to Suzuki (1987) because of leakage at edges of the flow restricting plates. Furthermore, this is not possible with U.S. Pat. Nos. 3,942,552 or 3,939,868 to Logsdon (1976) because the counterbalance spring never totally releases the flow restricting means or with U.S. Pat. No. 4,009,826 to Walker (1977) and U.S. Pat. No. 3,204,664 to Gorchev (1965) because of leakage at the edges of their sliding flow restrictors.
F—PULSATION
A phenomena that is well known in prior art and unique to “airflow powered” air volume regulators is their propensity to flutter, oscillate or pulsate when the air stream at their inlet is unstable. This inherent characteristic is due to the use of a spring to counterbalance the airflow restricting forces within the air volume regulator. Variations in the pressure upstream from the air volume regulator caused by turbulence or other instabilities can induce pressure pulses that travel down the ductwork to the air volume regulator. These fluctuations induce a rapid rise and fall in pressure usually lasting less than a second. If the amplitude of the pressure pulse is significant, the air volume regulator will react rapidly to constrict the airflow passage on sensing the rise in pressure then open the airflow passage on sensing the drop in pressure. But the inertia of the apparatus is such that the air volume regulator will tend to be out of phase with the quick change in pressure: over-constricting the airflow passage as the pressure starts to return to normal or under-constricting the airflow passage once the pressure has returned to its initial level. This out of phased reaction sets in motion the pulsation, as the spring-inertia combination oscillates between extremes driven by the energy of the air upstream of the apparatus as a car with defective shock absorbers when it hits a bump in the road. Dampening means must then be included to brake the cycle.
In prior art, U.S. Pat. No. 3,276,480 Kennedy (1966) and U.S. Pat. No. 3,763,884 to Grassi et al. (1973) employ dashpots and U.S. Pat. No. 3,049,146 to Hayes proposes a wear plate but their inherent friction hinders the airflow tracking of the air volume regulator and creates an unacceptably large hysteresis in its control. It is to be noted again that all mechanical friction within the apparatus prevents it from operating at low pressures. U.S. Pat. No. 3,204,664 to Gorchev (1965) teaches an air bellows with a flow orifice but the entrapped air is compressible and acts like a spring (air spring). The addition of mass to create inertia such as flywheels is shown in U.S. Pat. No. 3,060,960 to Waterfill (1962), but this method only lowers the natural frequency of the spring-mass combination: pulsation can still occur but at a lower frequency. The only true dampening that will dissipate the energy is due to the mechanical friction of this device. Under certain condition, I have found that the addition of inertia alone is ineffective.
In summary, the major drawbacks in prior art “airflow powered” air volume regulators are the following:                The minimum static pressure required by the air volume regulator to start controlling the airflow rate is relatively high. The long-felt need for an air volume regulator functioning reliably at pressures at or below 25 pascals (0.1″ w.g.) is unsolved.        The means for varying the airflow rate setpoint remain relatively complex, ineffective or in some cases, none existent. In some prior art embodiments, the flow rate set point may “slide” when a pneumatic actuator is employed.        Most prior art embodiments cannot attain “zero flow” when an actuator is proposed to vary their flow rate setpoint.        They have a propensity to flutter, oscillate or pulsate when the airstream at their inlet is unstable. Dampening means are proposed but either generate excessive mechanical friction or, under certain conditions, are ineffective.        